Self-pressurizing seal for rotary shafts

ABSTRACT

A self-pressurizing shaft seal for an oil filled submersible motor is characterized by an inboard spiral grooved face seal and an outboard conventional face seal disposed in series relationship along the motor shaft. During operation, the inboard spiral grooved face seal pumps oil from the motor interior into a substantially confined zone between the seals to increase the oil pressure at the outboard face seal without the necessity for structurally strengthening the entire motor housing. Also disclosed is the disposition of a shoulder on the spiral grooved seal runner face remote from the pumping interface to permit the oil pressure within the confined zone to hydraulicly increase the axial force upon the runner thereby increasing the pumping pressure of the spiral grooved seal in boot strap fashion. Other disclosed seals contain means for measuring the pressure within the confined zone to actuate remote signaling devices upon a failure of the outboard seal as well as spiral grooved face seals having valving means to alter the pumping rate of the inboard seal upon a loss of pressure in the confined zone between seals.

United States Patent MeHugh 1 Nov. 6, 1973 SELF-PRESSURIZING SEAL FORROTARY Primary ExaminerSamuel B. Rothberg SHAFTS Att0rneyFrank L.Neuhauser, Oscar B. Waddell, Jo- [75] Inventor: James Dennis McHugh,Santa Clara, Seph Forman and Van Myles Calif. [57] ABSTRACT [73]Asslgnee: 2 s Ycompany A self-pressurizing shaft seal for an oil filledsubmersc enecta ible motor is characterized by an inboard spiral [22]Filed: Apr. 8, 1971 grooved face seal and an outboard conventional faceseal disposed in series relationship along the motor [2}] App! l32485shaft. During operation, the inboard spiral grooved face seal pumps oilfrom the motor interior into a sub- [52] U.S. CI 277/2, 277/61, 277/65stantially confined zone between the seals to increase [51] F16j 9/00,F16j 15/38 the oil pressure at the outboard face seal without the [58]Field of Search 277/2, 3, 61, 65 necessity for structurallystrengthening the entire motor housing. Also disclosed is thedisposition of a [56] References Cited shoulder on the spiral groovedseal runner face remote UNITED STATES PATENTS from the pumping interfaceto permit the oil pressure 3 675 935 7/1972 Ludwi 277,25 within theconfined zone to hydraulicly increase the 3499653 3/1970 Gard":"2"77/96x axial force upon the runner thereby increasing the 2:628:8522,1953 voytech: 277ml x pumping pressure of the spiral grooved seal inboot 3 75 933 972 Nappe 277/65 strap fashion. Other disclosed sealscontain means for 3,176,996 4/1965 Barnett 277/3 X measuring thepressure within the confined zone to ac- 3,587,405 6/1971 Holmes 277/2tuate remote signaling devices upon a failure of the out- 3,392,9337/1968 j 277/2 X board seal as well as spiral grooved face seals having3339-930 9/1967 Tracy 277/65 X valving means to alter the pumping rateof the inboard gi f" seal upon a loss of pressure in the confined zonebe- 1C0 a1 1,989,548 1 1935 Coberly 277/2 x twee seals 8 Claims, 7Drawing Figures a M .7 [K Y 52 56 54 5o v s %n1um%% 4/ n eeqznnn ssm 44m1 2 1- 1 r r PATENTEU NEW 8 I973 SHEET 3 BF 4 PATENTEUHHY 619753770.179

SHEET w 4 FIG? SELF-PRESSURIZING SEAL FOR ROTARY SHAFTS This inventionrelates to a shaft seal for a rotatable machine and, more particularly,to a shaft seal wherein a high pressure zone of sealing fluid is formedbetween a spiral grooved face seal and a conventional face seal toinhibit ingress of contaminating fluid into the machine.

One of the major factors limiting the life of submersible motors iswater in-pumping at the shaft seal produced by a slight eccentricity inthe face seal customarily employed to assure minimum leakage at theshaft. Although water in-pumping can be overcome by substantiallyincreasing the pressure differential between the sealing fluid, e.g.,oil, typically contained within the motor and the ambient water, higherpressure differentials necessarily require reinforcement of the motorhousing as well as substantial alterations in the spring biaseddiaphragm customarily utilized to produce the oil/water pressuredifferential.

Because of the difficulties associated with increaisng the oil pressurewithin the motor, a number of different seal configurations have beenproposed to inhibit inpumping notwithstanding a low oil/water pressuredifferential. For example, rotor shafts have been sealed utilizing anexternal pump to produce high and low pressures within sealing chamberssituated at axially displaced locations along the shaft. Similarly, ithas heretofore been proposed that the shaft of a centrifugal pump besealed utilizing the rotary speed of the shaft to pump oil from anaxially outboard location to an inboard seal to restrict gas leakagefrom the pump. 1 also have proposed in my co-pending U.S. Pat. No.3,704,0l9, issued Nov. 28, 1972, and assigned to the assignee of thepresent invention utilization of a spiral grooved face seal having deephelical grooves to increase the pressure at the seal interface withoutincreasing the outpumping rate of oil from the motor. While all thesedesigns have certain advantages, there still remains a need for seals ofdifferent designs with differing capabilities.

It is therefore an object of this invention to provide a novelself-pressurizing seal characterized by low leakage.

It is also an object of this invention to provide a seal adaptable tomonitoring at an external location to assure proper seal functioning.

It is a further object of this invention to provide a self-pressurizingseal wherein the pressure of the sealing liquid within the seal isemployed to augment the mechanical bias of the seal thereby maximizingthe obtainable pressure from the seal without extensive wear of the sealduring start-up.

It is a still further object of this invention to provide aself-pressurizing seal wherein automatic closure of the seal is effectedupon a reduction in seal pressure.

These and other objects of this invention generally are achieved by aself-pressurizing' seal for a rotatable machine characterized by aninboard pumping seal having an annular running member mounted upon arotatable shaft in juxtaposition with an annular co-planar stationarymember. At least one of the juxtaposed members is provided with spiralgrooves extending from the perimeter of the member to a land along theplanar face of the member to pump sealing fluid from the rotatablemachine into a substantially confined zone. The pumping action of theinboard seal increases the pressure of the sealing fluid within the zonerelative to the sealing fluid pressure within the machine andconventional face seal means are disposed along the shaft at an axiallyoutboard location (relative to the inboard pumping seal) to restrict theflow of sealing fluid from the high pressure zone into the ambientwater. Because sealing fluid at relatively high pressure is situatedonly within a zone intermediate the axially displaced face seals,ingress of water into the motor is inhibited without structuralreinforcement of the entire motor housing and without subjecting thenecessary flexible oil expansion system to large pressure differences.

Although the features of this invention are defined with particularityin the appended claims, a more complete understanding of the inventionmay be obtained from the following detailed description of variousspecific embodiments when taken in conjunction with the appendeddrawings therein:

FIG. 1 is an enlarged sectional view of a selfpressurizing seal inaccordance with this invention,

FIG. 2 is a plan view of one member of the inboard face sealillustrating the disposition of spiral grooves therein,

FIG. 3 is a sectional view of a self-pressurizing seal wherein thehydraulic pressure of the oil within the seal is employed to increasethe obtainable seal pressure,

FIG. 4 is an enlarged sectional view illustrating the force distributionalong the spiral grooved seal of FIG.

FIG. 5 is an alternate seal configuration illustrating a seal monitoringdevice in accordance with this invention,

FIG. 6 is a sectional view of a spiral grooved seal wherein theoutpumping rate of the seal is reduced upon a reduction in the outboardseal pressure, and

FIG. 7 is a plan view of the stationary member forming the seal of FIG.6.

A self-pressurizing seal 10 in accordance with this invention isillustrated in FIG. I and generally includes an inboard spiral groovedface seal 12 and an outboard face seal 14 disposed in tandem upon shaft16 of a dynamoelectric machine, e.g., the pump motor such as isdescribed in U.S. Pat. No. 2,790,916, issued Apr. 30, 1957 to MB.l-Iinman (the entire disclosure of which patent is incorporated hereinbyreference). Typically, the pump motor contains a sealing fluid, e.g.,transformer oil 20, biased by a flexible diaphragm to increase thepressure of the oil approximately 5 psi relative to the water 24 whichforms the ambient environment for the motor during operation. The oilwithin the motor is in communication with the radially outer surface ofinboard seal 12 and is pumped by the spiral grooves of the inboard faceseal into substantially closed annular oil chamber 26 thereby increasingthe oil pressure of the chamber relative to the oil pressure within thepump motor.

Spiral grooved face seal 12 generally is characterized by an annularcarbon runner 28 mounted upon shaft 16 with planar face 30 of the runnerbeing disposed in a confronting attitude with planar face 32 of ceramicstationary member 34 fixedly secured to pump motor housing 36. One ofthe planar faces of face seal 12, illustrated in FIG. I as face 32 ofstationary member 34, has spiral grooves 22 therein to pump oil from themotor upon rotation of runner 28 relative to stationary member 34. Thegrooves, shown more clearly in FIG.

2, have a geometric configuration and density dependent upon thequantity of pumping desired by the face seal (as will be more fullyexplained hereinafter) and desirably extend radially from the outercircumferential edge of annular stationary member 34 to an annular land38 separating the grooves from central aperture 40 extending axiallythrough the member. In the event a failure of outboard face seal 14should necessitate a shutdown of the motor, land 38 advantageouslyfunctions to block back flow of water through spiral grooved face seal12.

Returning again to FIG. 1, the face of carbon runner 28 remote fromplanar face 30 is notched to form a lower shoulder 42 which, inassociation with backing plate 44, serves to house O-Ring 46 sealing thecarbon runner to shaft 16. A second shoulder 48 also is formed along theradially outer face of carbon runner 28 to seat a generally L-shapedbrass ferrule 50, biased against the runner by spring 52. To permitaxial movement of the ferrule along shaft 16 while restricting movementof the ferrule in a plane perpendicular to the shaft, elongated body 54of the ferrule is slidably engaged within a guide 56 fixedly secured tothe motor shaft.

Outboard face seal 14 is conventional in design and is mounted in tandemwith spiral grooved face seal 12 so that the pressure of the oil withinchamber 26 tends to close carbon runner 56 upon confronting ceramicstationary member 58. A biasing spring 60 augments the oil pressuretending to close the face seal by providing an axial force against upperextension 62 of ferrule 64 to drive inwardly extending backing plate 66toward carbon runner 56. The edge of ferrule 64 proximate spiral groovedface seal 12 extends through guide 68 to limit the axial movement of theferrule while sealing of the runner to the shaft is accomplished by aflexible bellows 70 fixedly secured between the shaft and the overlyingferrule.

To inhibit ingress of solid contamination into the motor, a sand slinger71 is secured to motor shaft 16 at an axial location to shroud theradially outer edge of outboard seal carrier 73. The seal carrier isfixedly mounted to the motor housing 36 by bolts 74 passing throughsuitable apertures in the outer flange of the seal carrier while aradially inner notch in the seal carrier serves as a seat for stationarymember 58 of face seal 14.

During operation of the motor, the rotary motion of carbon runnerrelative to spiral grooved stationary member 34 pumps oil from the motorhousing into annular oil chamber 26 to increase the oil pressure withinthe chamber to a predetermined level dependent primarily upon theanticipated water inpumping force at outboard face seal 14 resultingfrom eccentricity in the outboard seal. This predetermined pressurelevel can be calculated (in accordance with the teachings of an articleentitled Inward Pumping in Mechanical Face Seals, by .LA. Findlay,presented as paper No. 68 at the Lub 2 ASME-ASIE Lubrication Conference,Atlantic City, N.J., Oct. 1-10, 1968,) from the formula:

wherein Ap/e is the required oil pressure in lbs/in. for each incheccentricity (e) of outboard face seal 14,

e is the maximum tilt contemplated for face seal l4,

w is the shaft speed in radians per second,

R -R, is the radial span of the juxtaposed faces forming seal 14 ininches,

cosa is the maximum misalignment contemplated for face seal 14,

a is the viscosity of the water presumed to penetrate the sealinterface, in lb-sec./in. and

h is the average oil film thickness between faces of the seal in inches.Typically, an oil pressure increase of approximately 4,000 lbs/sq. in.is required to compensate for each inch of shaft eccentricity to assurezero inpumping at the outboard face seal.

Although the seal eccentricity can vary dependent upon such factors asthe amount of shaft runout under load and speed, the out-of-roundness ofthe carbon washer, etc., the total eccentricity generally can beestimated with a high degree of reliability for any given manufacturingprocedure. Thus, if manufacturing experience has indicated that aneccentricity of approximately 0.010 inch normally is not exceeded onfabricated face seals, the pressure required for chamber 26 to preventwater inpumping is calculated by multiplying the maximum observedeccentricity by the pressure per inch of face seal eccentricity ascalculated by the foregoing Findlay equation, e.g., for an empiricallydetermined maximum eccentricity of approximatley 0.010 inch and acalculated oil pressure of 4,000 psi per inch eccentricity, a totalpressure of 40 psi is required in oil chamber 26 to inhibit inpumping.

The outpumping rate at outboard face seal 14 also must be considered toassure that the oil supply within the motor is not exhausted within anunduly short time in an attempt to inhibit water ingress through theface seal. The outpumping rate for the outboard seal therefore iscalculated, e.g., from the formula:

wherein q is the outpumping rate,

Ap is the difference in pressure across face seal 14 in ps1,

11 is the average film thickness between juxtaposed faces of the seal ininches,

R; is the radius to the inner edge of the sealing land,

,u. is the viscosity of oil in the seal interface in lb.sec.- /in. and

AR is equal to the radial span of the juxtaposed faces forming the sealin inches. The optimum pressure for oil chamber 26 then is chosen as acompromise between the high oil pressure desired to overcome inpumpingof water into the motor and the low oil pressure desired to limit theoil outpumping rate at the outboard face seal.

Once the pressure desired for annular oil chamber 26 has been chosen,the geometric configuration of inboard spiral groove face seal 12required to produce this pressure can be determined in accordance withthe teachings of E.A. Muijderman in an article entitled SPl- RAL GROOVEBEARINGS published 1966 by Philips Technical Library. One spiral groovedface seal configuration found suitable for a [2 inch submersible motorhaving a 2 inch rotatable shaft was characterized by 10 equally spacedgrooves notched to a depth of 0.0013 inch and extending at a spiralangle of 15 with a groove land to width ratio of l. The inner and outerdiameters of the seal measured 1.87 inches and 2.37 inches, re-

spectively, while the groove inner diameter measured 1.95 inches. Withthe foregoing seal rotating at a speed of 30 revolutions per second, amaximum pressure of 67.5 psi was observed in annular oil chamber 26.

FIG. 3 illustrates an improved embodiment of this invention whereby theforce of the spiral grooved face seal biasing spring can be reducedwithout a reduction in the pressure obtainable from the face seal. Toachieve this result, a shoulder 72 is notched in carbon runner 28A at anoutboard location relative to O-Ring 46A thereby permitting pressurizedoil within annular oil chamber 26A to communicate with face 75 andhydraulically drive runner 28A axially towards mutual contact withstationary member MA as the pressure within the oil chamber increases.Although a shoulder 76 is provided in shaft ll6A to seal notched runner28A and the position of the back support for spring 60A has been changedslightly, the self-pressurizing face seal of FIG. 3 otherwise issubstantially identical to the face seal illustrated in FIG. 1. Theincreased axial force upon runner 28A, however, resulting from hydraulicpressure on face 75 reduces the gap of the spiral grooved face sealtending to increase the obtainable pressure from the seal. Thisincreased pressure, in turn, results in an increased hydraulic forceupon face 75 and the pressure within annular oil chamber 26A isincreased in bootstrap fashion until an equilibrium pressure is reached.

Assuming zero net flow at outboard face seal 14A, the pressure generatedby spiral grooved face seal 12A (illustrated by pressure diagram P inFIG. 4) increases approximately linearly from the outer periphery ofstationary member 34A to the inner extent of the grooves in thestationary member, i.e., from d to (1 with the pressure along theungrooved portion of the seal interface, i.e., from d to d,, remainingconstant at P For simplicity, the average pressure acting over the areabetween d2 and d may be assumed equal to l/2P The maximum pressure atequilibrium therefore can be estimated from the approximate formula:

qtma-m 2 wherein F, is the axial load upon the seal produced by spring52A in pounds,

d is the internal diameter of the spiral grooved annular portion of theface seal,

d is the external diameter of the spiral grooved annular portion of theface seal, and

d. is the diameter of hydraulic shoulder 72 formed in carbon runner 28A.One bootstrap seal having a seal inner diameter (i.e., d of 1.87 inches,a groove inner diameter (i.e., d of 1.95 inches, a seal outside diameter(i.e., d of 2.374 inches and a seal balance diameter (i.e., d of 2.0inches produced a hydraulic load of 28.7 lbs. upon the face seal inaddition to a bias of 42 lbs. provided by spring 52A for a total faceseal axial load of approximately 70.7 lbs. The spiral grooved runner ofthe face seal contained equally spaced grooves disposed at a spiralangle of 15 and the runner was rotated at a speed of approximately 30revolutions per second.

When the required pressure for the intermediate oil chamber is low,e.g., approximately 20 psi, the outboard face seal can be disposed in aback-to-back configuration with the inboard spiral grooved face seal asillustrated in FIG. 5. The pressure within oil chamber 26B then appliesan axial force upon outboard face seal 14B tending to separate theconfronting faces of the seal requiring a biasing spring 603 having anaxial force sufficient to overcome the hydraulic pressure within chamber2613 to maintain the desired outboard face seal opening duringoperation. If a failure of pressure should occur within chamber 268, thehydraulic force tending to maintain the outboard seal open would beremoved and biasing spring 603 would tend to close the faces of theoutboard seal inhibiting ingress of water into the motor. When theback-toback seal arrangement is utilized with relatively high sealpressures, e.g., pressures of approximately psi, care must be taken tochoose a biasing spring 60B having sufficient force to inhibit excessiveoutpumping of oil through the outboard face seal.

A major feature of this invention is the ability to monitor sealoperation at an external location by the disposition of a pressuretransducer 77 within oil chamber 26B as illustrated in FIG. 5. Thepressure transducer is connected in series with an alarm 78 and avoltage source, e.g., a transformer 80 having a primary winding 80Aconnected across the motor energization leads (not shown), and functionsto close the series circuit upon a reduction in pressure within oilchamber 26B below a predetermined minimum. Alarm 78 then is soundedpermitting shutdown and removal of the motor from a submerged locationprior to permanent damage of the motor interior by water seepagetherein. Should the pressure drop in chamber 26B be produced by afailure of outboard face seal 14B, seepage of water through spiralgrooved face seal 123 during shutdown is inhibited by annular land 38 ofthe face seal. To effectively function as a flow restricter during motorshutdown resulting from failure of outboard seal 1413, the annular landdesirably should have a radial span of at least 0.04 inches.

A self-contained motor protective device is illustrated in FIGS. 6 and 7wherein a spring loaded pressure relief valve 82 is employed to alterthe operation of inboard spiral grooved seal 12C from a full film to asolid-solid contacting mode in the event of failure of the outboardseal. Relief valve 82 functions to restrict the flow of oil from anannular groove 84 situated at the radially inner terminus of the spiralgrooves 22C to a bypass port 86 during normal operation of theselfpressurizing seal. If the outboard face seal should fail duringmotor operation reducing the pressure within oil chamber 26C confinedbetween the face seals, the hydraulic pressure on piston 87 of valve 82communicated to the valve through axial aperture 88 also drops and therelatively higher pressure of the oil within annular groove 84 overcomesthe bias of spring 90 to relieve the pressure at the seal interfacethrough bypass port 86.

With valve 82 open, the operation of spiral grooved face seal 12C shiftsfrom a conventional thick film operation, i.e., a film in excess ofapproximatley microinches typically produced by a conventional groovedepth of 1,000 to 1,500 microinches, to a solid-solid contacting mode,i.e., a film width below approximately 50 microinches, substantiallylimiting the outpumping rate of the spiral grooved face seal. Thus, aportion of the oil pumped by the spiral grooved face seal is valved backto the suction side of the face seal thereby reducing both the maximumpressure generated between faces of the spiral grooved face seal and thequantity of oil pumped into oil chamber 26C.

It will be appreciated that the spiral groove face seal will tend toclose, even without operation of relief valve 82, upon failure of theoutboard seal because of increased maximum pressure at the spiral grooveseal interface resulting in a changed oil distribution at the sealinterface. If the maximum pressure required by the seal under conditionsof leakage exceeds the maximum generating capacity of the seal, the sealwill inherently change from a full film mode to a solid-solid contactmode to reduce the outpumping rate. Thus, by careful choice of spiralgroove design, e.g., spiral groove width, depth and length, a seal canbe fabricated wherein the desired pressure will be produced withoutboard seal 14C functioning properly in a full film mode while asubstantially reduced outpumping rate is produced upon failure of theoutboard seal.

The previously cited formula (3) for estimating the maximum pressurerise clearly illustrates the effect of the shoulder d of FIG. 4 upon thepressure created. Formula (3) assumed an average pressure P /Z over thatportion of the seal interface where the pressure changes. An alternate,theoretically exact formula for calculating the maximum pressure risemay be obtained by integrating the assumed linear pressure rise over thearea between diameters d and d of FIG. 4. If the seal diameter d, isequal to the diameter d no shaft shoulder exists and the formula forcalculating pressure rise becomes:

wherein P X is the maximum pressure generated by the seal in psi in afull film mode with zero leakage,

F, is the total force applied to the seal by the biasing spring inpounds,

d is the span from the radially inner face of the face seal to the shaftaxis,

:2 is the span from the radially inward end of the spiral grooves to theshaft axis, and

at is the span from the radially outer periphery of the spiral groovesto the shaft axis,

The ratio of the maximum pressure developed at the pumping seal withleakage at outboard seal 14 interface relative to the maximum pressurecapable of being developed by the seal with zero leakage then can becalculated from the formula:

P no leakage fi d3 1 3 d3 v From this ratio, the maximum pressurecapable of being developed by seal 12 with no restriction in leakage atthe outboard seal can be calculated to provide an indication of filmthickness arising from the pressure increase. When the calculatedmaximum film pressure under leakage conditions exceeds the maximumgenerating capability of the seal (as can be calculated from theheretofore cited Muijderman publication), the seal operation changesfrom a full film mode to a solid-solid contact mode upon failure of theoutboard seal.

It should be appreciated that very shallow (e.g., 50 microinches) orvery deep (e.g., 20,000 microinches as described in my heretofore citedpatent application, Ser. No. 47,824) grooves can be utilized for theinboard face seal to reduce outpumping upon failure of the outboardseal. However, because the pressures produced by these face seals duringnormal operation is difficult to predict due to variations in fluidviscosity at the seal interface, such seals generally are notrecommended for the inboard face seal.

What I claim as new and desire to secure by Letters Patent of the UnitedStates of America is:

1. A shaft seal for a rotatable machine to inhibit ingress of ambientfluid into said machine, said seal comprising an inboard pumping sealcharacterized by an annular running member mounted upon said rotatableshaft in juxtaposition with an annular stationary member disposed in aconfronting attitude relative to said running member, at least one ofsaid juxtaposed members having viscosity grooves therein extending froma peripheral edge of said member and terminating in a land along aplanar face of the member to pump sealing fluid contained within saidmachine into a relatively confined zone to substantially increase thepressure of said sealing fluid within said zone relative to the sealingfluid pressure within said machine, and face seal means axially mountedupon said shaft at an outboard location relative to said inboard seal torestrict the out flow of sealing fluid from said confined zone, saidface seal means including a rotorary member axially mounted upon saidrotatable shaft, a stationary member juxtaposed in a co-planar attituderelative to said rotary member, and means including a pressure actuatedmember for driving said rotary and stationary members toward mutualcontact.

2. A seal having a rotatable shaft according to claim 1 wherein saidinboard pumping seal and said face seal means are disposed in tandemalong said shaft, the pressure of said sealing fluid in said confinedzone tending to reduce the span between the stationary and rotarymembers of said face seal means.

3. A seal having a rotatable shaft according to claim 1 wherein saidinboard pumping seal and said face seal means are disposed inback-to-back configuration along said shaft, the pressure of saidsealing fluid in said confined zone tending to increase the span betweenthe stationary and rotary members of said outboard face seal means.

4. A seal for a rotatable shaft according to claim 1 wherein one memberof said inboard pumping seal is axially slidable along said shaft andfurther including a shoulder notched within the face of said axiallyslidable member situated remote from the stationary member forming saidseal, said shoulder being in communication with said sealing fluidwithin said high pressure zone to provide a hydraulic force tending tobias said annular running and annular stationary members into mutualcontact.

5. A sea] for a rotatable shaft according to claim 4 further includingan O-Ring seal between said rotary member and said shaft to seal saidrotary member upon said shaft.

6. A seal having a rotatable shaft according to claim 1 furtherincluding means for measuring the pressure within said confined zone andremote signal means responsive to said pressure measuring means forindicating a reduction in the pressure of said confined zone below apredetermined minimum.

7. A shaft seal for a rotatable machine to inhibit ingress of ambientfluid into said machine, said seal comprising a spiral grooved face sealdisposed at an inboard location along said shaft for pumping sealingfluid from said machine into a substantially confined zone to increasethe sealing fluid pressure within said machine, said spiral grooved faceseal comprising coaxial rotary and stationary members juxtaposed in aco-planar attitude, at least one of said members being axially slidablealong said shaft to vary the span between said members, mechanical meansbiasing said axially slidable member toward said stationary member, ashoulder notched within said axially slidable member face remote fromsaid stationary member, said shoulder being in communication with thesealing fluid of said confined zone to bias said axially slidable membertowards said stationary member of said spiral grooved face seal withincreased sealing fluid pressure in said confined zone, and pressureresponsive face seal means disposed at an axially outboard location uponsaid rotatable shaft and operable in response to increased fluidpressure in said confined zone to restrict the flow of sealing fluidfrom said confined zone.

8. A seal for a rotatable shaft according to claim 7 wherein the groovesin said spiral grooved face seal extend from the periphery of said sealto terminate in an annular land.

1. A shaft seal for a rotatable machine to inhibit ingress of ambientfluid into said machine, said seal comprising an inboard pumping sealcharacterized by an annular running member mounted upon said rotatableshaft in juxtaposition with an annular stationary member disposed in aconfronting attitude relative to said running member, at least one ofsaid juxtaposed members having viscosity grooves therein extending froma peripheral edge of said member and terminating in a land along aplanar face of the member to pump sealing fluid contained within saidmachine into a relatively confined zone to substantially increase thepressure of said sealing fluid within said zone relative to the sealingfluid pressure within said machine, and face seal means axially mountedupon said shaft at an outboard location relative to said inboard seal torestrict the out flow of sealing fluid from said confined zone, saidface seal means including a rotorary member axially mounted upon saidrotatable shaft, a stationary member juxtaposed in a co-planar attituderelative to said rotary member, and means including a pressure actuatedmember for driving said rotary and stationary members toward mutualcontact.
 2. A seal having a rotatable shaft according to claim 1 whereinsaid inboard pumping seal and said face seal means are disposed intandem along said shaft, the pressure of said sealing fluid in saidconfined zone tending to reduce the span between the stationary androtary members of said face seal means.
 3. A seal having a rotatableshaft according to claim 1 wherein said inboard pumping seal and saidface seal means are disposed in back-to-back configuration along saidshaft, the pressure of said sealing fluid in said confined zone tendingto increase the span between the stationary and rotary members of saidoutboard face seal means.
 4. A seal for a rotatable shaft according toclaim 1 wherein one member of said inboard pumping seal is axiallyslidable along said shaft and further including a shoulder notchedwithin the face of said axially slidable member situated remote from thestationary member forming said seal, said shoulder being incommunication with said sealing fluid within said high pressure zone toprovide a hydraulic force tending to bias said annular running andannular stationary members into mutual contact.
 5. A seal for arotatable shaft according to claim 4 further including an O-Ring sealbetween said rotary member and said shaft to seal said rotary memberupon said shaft.
 6. A seal having a rotatable shaft according to claim 1further including means for measuring the pressure within said confinedzone and remote signal means responsive to said pressure measuring meansfor indicating a reduction in the pressure of said confined zone below apredetermined minimum.
 7. A shaft seal for a rotatable machine toinhibit ingress of ambient fluid into said machine, said seal comprisinga spiral grooved face seal disposed at an inboard location along saidshaft for pumping sealing fluid from said machine into a substantiallyconfined zone to increase the sealing fluid pressure within saidmachine, said spiral grooved face seal comprising coaxial rotary andstationary members juxtaposed in a co-planar attitude, at least one ofsaid members being axially slidable along said shaft to vary the spanbetween said members, mechanical means biasing said axially slidablemember toward said stationary member, a shoulder notched within saidaxially slidable member face remote from said stationary member, saidshoulder being in communication with the sealing fluid of said confinedzone to bias said axially slidable member towards said stationary memberof said spiral grooved face seal with increased sealing fluid pressurein said confined zone, and pressure responsive face seal means disposedat an axially outboard location upon said rotatable shaft and operablein response to increased fluid pressure in said confined zone torestrict the flow of sealing fluid from said confined zone.
 8. A sealfor a rotatable shaft according to claim 7 wherein the grooves in saidspiral grooved face seal extend from the periphery of said seal toterminate in an annular land.